design and internal flow analysis of a ducted contra

DESIGN AND INTERNAL FLOW ANALYSIS OF A DUCTED CONTRA-ROTATING AXIAL FLOW FAN

Ali Mohammadi Amirkabir University of Technology

Tehran, Iran

Masoud Boroomand Amirkabir University of Technology

Tehran, Iran

ABSTRACT This paper presents the design procedure of a ducted contra-

rotating axial flow fan and investigates the flow behavior inside it

using ANSYS CFX-15 flow solver. This study investigates

parameters such as pressure ratio, inlet mass flow rate and

efficiency in different operating points. This system consists of

two rotors with an outer diameter of 434 mm and an inner

diameter of 260 mm which rotate contrary to each other with

independent nominal rotational speeds of 1300 rpm. Blades’

maximum thickness and rotational speeds of each rotor will be

altered as well as the axial distance between the two rotors to

investigate their effect on the overall performance of the system.

Designed to deliver a total pressure ratio of 1.005 and a mass

flow rate of 1.8 kg/s at nominal rotational speeds, this system

proves to be much more efficient compared to the conventional

rotor-stator fans. NACA-65 airfoils are used in this analysis

with the necessary adjustments at each section. Inverse design

method is used for the first rotor and geometrical constraints are

employed for the second one to have an axial inlet and outlet

flow without using any inlet or outlet guide vanes. Using free

vortex swirl distribution method, characteristic parameters and

the necessary data for 3D generation of this model are obtained.

The appropriate grid is generated using ATM method in

ANSYS TurboGrid and the model is simulated in CFX-15 flow

solver by employing k-ε turbulence model in the steady state

condition. Both design algorithm and simulation analysis

confirm the high anticipated efficiency for this system. The

accuracy of the design algorithm will be explored and the most

optimum operating points in different rotational speed ratios

and axial distances will be identified. By altering the outlet

static pressure of the system, the characteristic map is obtained.

INTRODUCTION There is currently a growing tendency toward developing

contra-rotating fans and compressor stages due to their

compactness in length and reduction in weight which is the result

of omitting stators from the conventional models. Many

researchers working in this area have reported higher quantities in

efficiency and pressure ratio for contra-rotating fans, propellers

and compressor stages in comparison to rotor-stator assembly.

Started in 1930s, Lesley [2] showed that an addition of fixed

counter-propeller blades increases the efficiency of a four-blade

fixed counter-propeller in combination with a two-blade rotating

propeller by two percent. In 1980s, Sharma et al. [3] examined

the effect of altering rotational speed ratios of two rotors and also

different axial spacing between them in one stage of a compressor

which had blades with 0.66 hub-tip ratio. It was concluded that

alteration of these factors has strong influence on the stalling

behavior of the stage. Also it was mentioned that if the second

rotor contra-rotates 50% faster than the first one, rotating stall

phenomenon would be suppressed and in case of large axial

distances, contra-rotation would lose its benefits. In 2009, an

increase in efficiency and a decrease in fuel consumption rate

were reported by Min et al. [4] in a contra rotating propeller

system. They claimed that by using the rotational flow energy

behind the first rotor, the propulsive efficiency of the system

would be improved. Later, Nouri et al. [5] further analyzed the

effect of the rotational speeds ratios of the rotors and the axial

distance between them in a low number of blade contra-rotating

axial flow fan. They reported optimum quantities for these

parameters and validated the results experimentally. Finally,

Mistry et al. [6] conducted a similar experiment on high aspect

ratio contra-rotating axial fan stage. They reported high quantities

of efficiency in design and off design operating points. In this

research, after choosing an optimum blade thickness, operation of

a high number of blades axial flow fan in different axial distances

and also with different rotational speed ratios would be

investigated to indicate its characteristic map. In near future, an

actual model will be built based on the design procedure

introduced in this article.

Proceedings of the ASME 2014 International Mechanical Engineering Congress and Exposition IMECE2014

November 14-20, 2014, Montreal, Quebec, Canada

IMECE2014-39883

1 Copyright © 2014 by ASME

NOMENCLATURE = Flow Angle in the Absolute Reference

= Flow Angle in the Relative Reference

' = Blade’s Angle

= Camber Angle

i = Incidence Angle

= Deviation Angle

= Stagger Angle

= Loading Coefficient

= Flow Coefficient

= Heat Capacity Ratio

Cp = Molar Specific Heat at Constant Pressure = Solidity = Density

Re = Reaction Ratio

T = Temperature

P = Pressure

N = Rotational Speed

R = Radius

RR = Rotational Speed Ratio

U = Linear Speed

DF = Diffusion Factor

R.P. = Required Power

L = Chord’s Axial Length

h = Blade’s Height

s = Axial Spacing

NoB = Number of Blades

A.R. = Aspect Ratio

A = Area

D = Diameter

P.R. = Pressure Ratio = Efficiency

m = Mechanical Efficiency (Engine)

= Loss Coefficient

= Loss Coefficient

m = Mass Flow Rate

Cx = Axial Flow Speed

w = Tangential Velocity

Cl = Lift Coefficient

Cd1 = Initial Drag Coefficient

Cda = Annulus Drag Coefficient

Cds = Secondary Drag Coefficient

tb/c = Thickness

Kt,i = Incidence Thickness Correction Factor

Kt,δ = Deviation Thickness Correction Factor

Ksh = Shape Factor

Subscripts:

1 = First Rotor

2 = Second Rotor

11 = Inlet of First Rotor

12 = Outlet of First Rotor

21 = Inlet of Second Rotor

22 = Outlet of Second Rotor

0 = Zero Camber Condition

10 = Parameter for 10% Thick Profile

o = Total Thermodynamic Condition

h = Blade’s Hub

m = Blade’s Mean-line

t = Blade’s Tip

DESIGN OF ROTORS Based on the fundamental concepts in Aungier [1], an inverse

design procedure is developed for the first rotor. The second rotor

is designed using some geometrical constraints between the outlet

flow angles of the first rotor and the inlet flow angles of the

second one. Simplified Radial Equilibrium is employed for this

procedure. Two 750 watt Electrical Motors with a maximum

rotational speed of 1385 rpm are assumed to be used in future

experimental analysis. Thus, the required power in the operating

points must not exceed this value. Setting the hub diameter (Dh) in

accordance with this engine diameter to be 26 cm and also

employing a hub-tip ratio of 0.598, the tip diameter (Dt) of these

contra-rotating rotors would be 43.5 cm. A 2.5 mm tip clearance

length is anticipated at the shroud section for safety consideration

in future experimental studies. Thus, the casing diameter (Dc)

would be 44 cm.

To have an axial inlet and outlet flow without using any Inlet

and Outlet Guide Vanes, the corresponding flow angles in the

absolute frame is set to be zero (Equation 1). According to the

selected electro-motor specifications, nominal design rotational

speeds are set to be 1300 rpm for both rotors. It is desired to have

an inlet mass flow rate of 1.8 kg and a total pressure ratio of

1.005. The mechanical efficiency of the selected electro-motor

(ηm) is 70% and the isentropic efficiency of the system (η) is

assumed to be 85%. Standard atmospheric condition is employed

(To = 288 K and Po = 101.3 kpa) which gives a density value of

1.226 kg/m3 in combination with equation 3 in an iterative

procedure. Equation 4 gives us the linear speed at each section

and equations 5 and 6 give us the required power for the first

rotor of this contra-rotating fan in the design rotational speed

which is equal to 624.326 watt considering the electro-motor’s

mechanical efficiency. Using equations 7 to 9, some

dimensionless variables such as the loading coefficient, the flow

coefficient and the reaction ratio are obtained. Table 1

summarizes the initial quantities obtained from equations 1 to 9.

01211 (1)

)(4

22

ht DDA

(2)

A

mCx

(3)


)(2 NRU (4)

]1..[

1

11

RPTT o

o

(5)

m

op TCmPR

2..

1

1

(6)

2

1

U

CT po

(7)

U

Cx

(8)

11tan2

1Re

(9)

Table 1 - Initial Parameters

Parameter Quantity

A (m2) 0.099

Cx (m/s) 14.835

Um (m/s) 23.596

ΔTo1(K) 0.241

R.P.1 (watt) 624.326

m1 0.436

m1 0.628

Re1m 0.782

As well as using the obtained dimensionless variables,

some geometrical constraints are employed between the first

rotor’s outlet flow angles and the second rotor’s inlet flow

angles in the absolute and also in the relative reference systems

to obtain other flow angles. Equation 11 states that the exiting

flow from the first rotor must enter the second rotor with a same

flow angle. The minus sign indicates the change of coordinate

system which is due to the contra-rotation of two rotors. The

velocity triangles in Figure 1 show the relation between these

flow angles.

)tan)Re1(2

(tan 11

1

11

12 m

m

mm

(10)

mm 2112 (11)

)tan1Re2

(tan 11

1

11

12 m

m

mm

(12)

)tanRe2

(tan 12

1

11

11 m

m

mm

(13)

)tan

(tan 12211

21

x

mxmmm

C

CUU

(14)

)(tan 21

22

x

mm

C

U

(15)

.

Figure 1 - Velocity Triangles of Rotors

To obtain the optimum solidity of blades for each rotor, the

procedure presented in [7] is employed as follows:

n

opt BA 2 (16)

where,

)(042231.00197.0 21 A (17)

))}(002368.033303.0(

)(427.13exp{

21

21

B

(18)

)(04677.08592.2 21 n (19)

Choosing a desirable aspect ratio of 2 for each rotor, the

number of blades and also the axial distance between them can

be determined using equations 20 to 22. Quantities obtained for

the number of blades are rounded to the nearest odd number.

Therefore, the first rotor will consist of 27 blades and the

second one consists of 21 blades. As the number of blades is

determined, axial spacing between them and also axial chord

lengths would be calculated again using equations 21 and 22.

Assuming the axial chord lengths to be constant in each section

and employing free vortex swirl distribution method, similar

calculation is done for all other sections of each rotor. Table 2

summarizes some of these obtained quantities. Then, equations

25 to 32 are used respectively to calculate the diffusion factor,

different drag coefficients, lift coefficient and loss coefficient.

Finally equation 33 gives the efficiency of the system. Further

explanations of these parameters are available in [1].


Table 2 - Geometrical Data of Both Rotors

hub

Rotor 1

Mid

Shroud

Hub

Rotor 2

Mid

Shroud

Inlet Relative Flow Angle (degrees) 50.012 57.581 63.298 64.721 66.273 68.533

Outlet Relative Flow Angle 15.012 41.156 55.097 50.027 57.581 63.298

Solidity 1.456 1.103 0.873 1.066 0.807 0.639

Incidence Angle 0.385 -0.792 -1.065 -2.688 -4.086 -5.678

Camber Angle 46.38 26.696 17.117 28.678 23.538 22.299

Deviation Angle 11.75 9.478 7.85 11.295 10.761 11.386

Stagger Angle 26.451 45.026 55.805 53.071 58.589 63.061

AR

hLm

(20)

mopt

mm

Ls

(21)

m

m

s

RNoB

2

(22)

tan xCw (23)

)tan(tan 21 xCw (24)

11

2

21

wL

sw

w

wDF

(25)

)tan(tan2

1tan 21 m

(26)

1

1

2cos

1w

(27)

mdL

sC 3

1 cos

(28)

mdm

l CL

sC

tan

)tan(tancos21

21

(29)

h

sCda 02.0

(30)

2018.0 lds CC (31)

dsdaddt CCCC 1 (32)

m

dtmt

CwL

3

1

cos

21 t

(33)

Aungier [1] mentioned that NACA-65 airfoil’s general

profile can be used to determine the coordinates of blades at

each desired section. This general set of coordinates is

available in Table 4-1 [1]. For this procedure, a zero camber

lift coefficient must be calculated. Using equations 34 to 45,

the incidence and the deviation angles are determined. Ksh is

set to be 1 for NACA-65 profiles [1]. As it will be introduced

further in the following section, maximum thickness of the

blades which is indicated by tb/c is assumed to be 0.08. This

means the first rotor blades are 3.528 mm thick and the second

rotor blades are 3.32 mm thick at their most.

q

bit ctK )/10(, (34)

])(1.0[

28.0

3.0

c

tq

b

(35)

]4/)70exp[(

1.0)3.2exp(465

)(

1

3110

*

0

p

i

(36)

160/914.0 3p (37)

43.05.1

)90/(06.0025.0

)2.11(

12n

(38)

210

*

0,

* )( niKKi itsh (39)

)09.167.1(

1

9.1

110

*

0

)90/(

]374.0[01.0)(

(40)

bmm /0.1 (41)

32

0.1 316.0132.0074.0249.0 xxxm (42)

385.017.09625.0 xxb (43)

2

, )/(5.37)/(25.6 ctctK bbt (44)

mKK tsh 10

*

0,

* )( (45)

To indicate the camber angles, equation 39 and 45 are

solved in an iterative manner. Alternatively, equations 46 to 49

may be employed directly by introducing A2 & B2 which are


derived from equations 34 to 45 and also from equations 50 to

53. Finally, equation 54 provides the zero camber lift

coefficient. Therefore, coordinates of blades at all sections are

now derived. Table 2 summarizes the obtained values at three

sections and Figure 1 summarizes the whole design process.

1

)()(

0.1

*

02

2

10

*

012

m

B

n

iA

(46)

1

)1(

0.1

0.1

2

22

m

m

n

nB

(47)

2

2'

11B

A

(48)

2

10

*

011

'

112

'

12

)()1(

n

in

(49)

'

12

'

111 (50)

'

11111 i (51)

'

12121 (52)

2

1'

111

(53)

4tan

1103.0

1 10

lC

(54)

Figure 2 - Summary of the Design Procedure

NUMERICAL MODEL Data obtained using the procedure explained in previous

section is used to generate the 3D model in ANSYS TurboGrid

which is also used for grid generation. ATM optimized

(Automatic Topology and Meshing) method is chosen over

other methods due to its better adoptability to blades surfaces.

Various grids with different number of elements are analyzed

for grid sensitivity analysis. The inlet total temperature and

pressure are set to be 288K and 101.3 kpa respectively while

the outlet boundary condition is determined using the outlet

static pressure. No slip wall condition is used for the hub and

blade sections and contra-rotating walls are used for the

shroud sections. Table 3 summarizes the results for grid

analysis while the outlet static pressure is 101.6 kpa. Also,

Figure 3 shows the generated model of the first rotor.

Table 3 - Grid Sensitivity Analysis for Contra-Rotating

Fan

1st Rotor

Surrounding

Nodes

2nd

Rotor

Surrounding

Nodes

Total

Pressure

Ratio

Total

Efficiency

52488 59236 1.0047 83.9877

129438 128840 1.0048 86.8611

241083 262742 1.0049 87.312

Figure 3 - Generated 3D Model for the First Rotor

The difference in total pressure ratio of the second and the

third sets of grid is less than 0.01% and the difference between

the efficiencies of these two sets is less than 0.52%. Thus, the

second set is selected to continue the analysis as using finer

grid shows no major improvement on the results. A gap equal

to 2.5 mm is anticipated between the blade’s tip and the

covering case to simulate the tip clearance length. Then the

model is simulated and solved using ANSYS CFX-15 flow

solver using standard atmospheric condition and also the

design parameters discussed before. k-ε turbulence model in


steady state condition is used as it is widely suggested in other

published researches in this area as in [6]. The final

computational grid consists of one single passage for each

rotor. The simulation is then generalized for the whole system

using periodic condition between the flow passages. The

second rotor’s domain starts exactly where the first domain

ends as it is shown in Figure 4. For cases with an axial distance

greater than 2 cm, a stationary part is added to the model to

increase the dignity of the solution. Figure 5 shows each of

these models.

Figure 4 - Generated Grid around the First Rotor (right)

and the Second Rotor (Left)

Figure 5 - Components in Different Axial Distances: a- 2cm

b- 4 to 12cm

For validation of the turbulence model, NASA Stage 37 [8]

was analysed in two different rotational speeds (70% and 90% of

the Nominal Rotational Speed). The reason that NASA stage 37

is preferred over other published models such as NASA rotor 67

for validation in this project is that this model is consisted of two

consecutive frames. Similarly, the current understudied project

consists of two frames in most cases. In 70% of nominal

rotational speed, the flow is mainly completely in the subsonic

flow regime; thus, it perfectly matches this subject. It should be

mentioned that data (geometry of the blades and simulation data)

for a more similar case which is consisted of two consecutive

rotors was not fully available to be used. Table 4 contains

information regarding the grid sensitivity analysis of this model

while operating with an outlet static pressure of 101.3 kpa. The

difference between the total pressure ratio in the second and the

third sets of generated grids is 0.015% and the difference

between the efficiencies is 0.064%. So the second set is chosen

for validation purposes as no major improvement is made in

results using finer grids.

Table 4 - Grid Sensitivity Analysis for NASA Stage 37

1st Rotor

Surrounding

Nodes

2nd

Rotor

Surrounding

Nodes

Total

Pressure

Ratio

Total

Efficiency

45710 37408 1.291 83.7326

169735 169101 1.2905 84.4847

323282 329529 1.2903 84.5393

Figure 6 shows a comparison between the experimental data

of NASA stage 37 report and the results obtained from

simulation of this model. Comparison of the pressure ratios in

both rotational speeds which is presented in Table 5 shows that

at its most, the results are less than 6% different which shows a

great agreement. As the validation of the turbulence model is of

interest here and the results show a similar trend and an

acceptable difference, k-ε will be used to continue the analysis.

Figure 6 - Comparison between the Simulation Results and

the Experimental Report of NASA Stage 37


Table 5 - Comparison of Results in Similar Inlet Mass Flow

Rates

Inlet Mass Flow Rate (kg/s) Difference in P.R. (%)

18.89 5.519

17.91 1.18

17.26 1.147

15.38 4.161

14.81 3.168

14.11 2.823

13.24 2.284

RESULTS AND DISCUSSION

Effect of Blade’s Thickness To investigate the effect of

blade’s thickness on the overall performance of the system, three

different quantities for blade thickness are chosen and used for

preliminary simulation which is presented in Table 6. For this

step, the outlet static pressure is chosen to be 101.65 kpa.

Table 6 – Effect of Various Airfoil Thicknesses

Max. Blade

Thickness (tb/c)

Inlet Flow

Rate (kg/s)

Pressure

Ratio

Efficiency

0.08 1.6778 1.0051 83.8157

0.1 1.6332 1.005 83.1111

0.12 1.5771 1.005 82.1774

Table 7 - Comparison of Parameters between Different

Blade Thicknesses and tb/c = 0.08

Max. Blade

Thickness

(tb/c)

M.F.R

Difference

% (kg/s)

P.R.

Difference

%

Efficiency

Difference

%

1 -2.658 -0.0099 -0.841

1.2 -6.001 -0.0099 -1.955

It is seen that thickening the blade profiles decreases the

possible inlet mass flow rate as it blocks some of the available

flow passage. Also, the total pressure ratios for the second and

third blade thicknesses are decreased compared to the one

related to maximum blade thickness of 0.08. The same pattern

occurs in the efficiency of the system. Table 7 shows the

comparison between the parameters obtained by the first blade

thickness (tb/c = 0.08) and the two other sets. Based on the

stated results, it is decided to choose 0.08 for the maximum

thickness chord ratio to further analyze the behavior of this

contra-rotating fan.

Accuracy of the Design Algorithm Results obtained from

simulation is used to verify the accuracy of the design algorithm.

Table 8 shows the consistency of results obtained from the

design algorithm and the simulation analysis. It is seen that the

total pressure ratio is slightly over predicted and the efficiency is

slightly under predicted in the design algorithm.

Table 8 - Validation of the Design Algorithm

Parameter Design

Algorithm

Simulation

Analysis

Difference

(%)

Inlet Flow Rate (kg/s) 1.8 1.8 0

Total P.R. 1.005 1.0046 0.039

Total Efficiency 85.0 86.4 1.647

Effect of Axial Distance Variation Six different axial

distances between two rotors are investigated starting from 2cm

(50% chord of the first rotor’s blade) to 12cm (300% chord of

the same blade). For the axial distance of 2cm, two rotating parts

are considered. For all other five cases from 4cm to 12cm, three

components are considered including one stationary part

between the rotors.

Figures 7a to 7c show the effect of variation in axial

distance on the operation of this contra-rotating system. Figure

7a shows that variation in axial distance has a negligible effect

on the inlet mass flow rate when the outlet static pressure is

altered. Figure 7b show that increasing this distance doesn’t

necessarily improve the pressure ratio of the system. Yet, Figure

7c shows that the CRF (Contra-Rotating Fan) operates with the

highest efficiency in the minimum axial distance. There can be

two reasons for this phenomenon. The first reason is that the

rotational energy produced by the first rotor can be best utilized

by the second rotor in case of minimum axial distance.

Increasing this distance results in dissipation of the rotational

energy behind the first rotor and thus decreases the efficiency of

the system. The second reason is that at this minimum axial

distance, the flow will be delivered to the second rotor at its

most precise quantity of the inlet flow angle. All other cases

from 4cm to 12cm show a very similar pattern in efficiency

while the inlet mass flow rate is altered.

Another interesting phenomenon is that when the system

is operating at the minimum anticipated length, the stall margin

would be widened. Figure 7a to 7c show that increasing this

distance from 2cm to 6cm limits the operating range. However,

between 6cm and 12cm, there seems to be no variation in this

issue. The maximum outlet static pressure which corresponds

to the stall limit is plotted vs. different axial distances in Figure

8. This can also be observed in relative Mach no. plots as well.

The velocity contours of the first rotor for all axial distances

are similar. Thus, only the one related to the minimum axial

distance is presented in Figure 9. Over the first rotor, the flow

is completely in low subsonic regime as anticipated.


Figure 7a-c Performance Parameters of CRF in Different

Axial Distances

Figure 8 - Stall Outlet Static Pressure vs. Axial Distance

Figure 9 - Relative Mach No. Contours for the First Rotor

in Axial Distance = 2cm

In the second rotor, there is a small change of relative

velocity in different axial distances. While operating at the

minimum distance, the blades experience a slightly less

relative Mach no. compared to other cases which is shown in

Figure 10a. Increasing the distance from 4cm to 12cm has no

major influence on this parameter. This is shown in Figures

10b and 10c where the Mach no. contours are similar. All

contours in Figure 9 and 10a to 10c are obtained when the

outlet static pressure is 101.6 kpa. Because of better efficiency

of the system in the minimum axial distance which is mentioned

above, this case is chosen to be used for investigation of the

effect of rotational speed ratio on this system.

Effect of Rotational Speed Ratio Introducing RR as the

rotational speed ratio, five different cases will be analysed.

1

2

N

NRR

(55)


Figures 10a-c Relative Mach No. Contours for the Second

Rotor in Axial Distances of a- 2cm, b- 4cm and c- 12cm

These five cases are: 1300 - 1150 (RR=0.884), 1300 -

1225 (RR=0.942), 1300 - 1300 (1.00), 1300 - 1385 (1.065)

and 1385 - 1300 (RR=0.938). Figure 10a to 10c show that

when N1 is fixed at 1300 rpm, increasing N2 improves the

performance of the system. Figure 11a shows that increasing

N2 results in increasing of the inlet mass flow rate while the

outlet static pressure is constant. Also Figures 11b and 11c

show that the system delivers higher pressure ratio and

operates with a better efficiency when N2 is increased. The

reason is the suction effect of the second rotor. Increasing N2

increases the suction effect of this rotor which eliminates the

wake produced by the first one. Comparing 1300 - 1225 set to

1385 - 1300 set which has nearly the same RR (0.94) shows

that the second set has better performance. This is because

increasing N1 to 1385rpm decreases the wake produced after

the first rotor and also the afterwards flow separation.

Figure 11a-b Performance Parameters of CRF in different

Rotational Speed Ratios


Figure 11c - Performance Parameters of CRF in different

Rotational Speed Ratios

Figures 12a to 12c show the velocity contours of the first

rotor operating in three different RR. It is seen that although

N2 has increased from 1150 rpm to 1385 rpm, there is no

major change in magnitude or direction of velocity vectors of

the first rotor while N1 is constant. Increasing N1 to 1385 rpm

results in higher magnitudes of velocity as it can be seen in

Figure 12c.

Figure 12a-c Velocity Contours for the First Rotor in

a- 1300-1150 set, b- 1300-1385 set and c- 1385-1300 set

Figures 13a to 13c show the velocity contours for the

second rotor. Comparing figure 13a and 13b shows that

increasing N2 has a little effect on the direction of the velocity

vectors. However it increases the velocity magnitude. On the

other hand, increasing N1 has a little effect on either the

direction and also the magnitude of velocity vectors of the

second rotor.


Figure 13a-c Velocity Contours for the Second Rotor in

a- 1300-1150 set, b- 1300-1385 set and c- 1385-1300 set

CONCLUSIONS A step by step design procedure of a contra-rotating axial

flow fan is presented. Then the 3D model is generated and the

appropriate grid is applied to the model in ANSYS Turbogird.

Finally the flow field around the rotors is simulated using CFX-

15 flow solver to determine characteristics of this system in

various operating conditions. Due to the results obtained, the

following conclusion can be mentioned:

1. Airfoil thickness has an important influence on the

efficiency and pressure ratio of the system. It is decided to

choose NACA-65 airfoil profiles with a maximum

thickness to chord ratio of 0.08 due to its higher inlet mass

flow rate in the same operating condition compared to

other cases and also because of its higher efficiency.

2. Analysis of different axial distance between the rotors

show that an axial distance equal to 50% of chord is the

most optimum distance for this CRF. It is shown that

employing this axial distance results in obtaining higher

efficiency and also a wider operating range which delays

the stalling condition. Further increase in the axial length

between two rotors, from 6cm to 12cm shows no

significant influence on the performance of the system.

3. Analysis of the rotational speed ratios (RR) indicates

that as the rotational speed ratio increases, better

performance is obtained from this system in terms of

efficiency and pressure ratio. This is because of the

stronger suction effect imposed by the second rotor on the

first one as N2 increases. Also, increasing N1 reduces the

produced wake by this rotor which results in better

performance. The best performance is achieved while the

system is operating with a RR ratio of 1.065 and an axial

distance equal to 50% of the first blade’s axial chord.

Based on the results obtained, an experimental test stand for

this contra-rotating axial flow fan will be built at DANA

Research Laboratory. As the experimental data is not available at

this time, a previous published case of NASA is used for

validation of the turbulence model and results of the simulation

analysis. Once the system is constructed, experimental data will

be compared to what obtained from simulation. Further

computational study is also underway to explore the effect of

different tip clearances and also higher rotational speed sets.

REFERENCES 1. Aungier, R.H. [2003], Axial Flow Compressors, A strategy

for Aerodynamic and Design Analysis Book, ASME Press.

2. Lesley, E. [1993], Experiments with a Counter-Propeller

Tech. Rep. 453, National Advisory Committee for

Aeronautics, Washington, USA.

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